Internal combustion engine variable throw crankshaft

ABSTRACT

A crankshaft for an internal combustion engine comprising a plurality of crank throws whose crank radii are made to vary throughout the crankshaft revolution. The crankshaft and varying radii crank throws are interconnected by four gears which rotate an eccentric which is placed between the connecting rod and crankpin. Rotation of the eccentric may be clock-wise or counterclockwise depending on the arrangement of gears used to generate either epicycloidal or hypocycloidal motion in which the cuspated motion is attenuated by incorporating a pair of noncircular gears which also rotate the eccentric upon the crankpin in a manner which causes the piston to move in the cylinder to a precalculated position during the combustion independent of the uniform speed of the crankpin.

United States Patent Mcwhorter [54] INTERNAL COMBUSTION ENGINE VARIABLETHROW CRANKSHAFI [72] Inventor: Edward M. Mcwhorter, 6931 GreenbrookCircle, Citrus Heights, Calif. 95610 [22] Filed: May 28, 1970 [21] Appl.No.: 41,298

[56] References Cited UNITED STATES PATENTS 5/1907 Cole 123/78 F X5/1940 Fiala-Fembrugg 123/78 F X 2/1945 Muml X us: 3,686,972 an Aug. 291912 UPI-1E1! PUBLICATIONS U.S. Dept. of Commerce, General InformationCon cerning Patents.

Primary Examiner-William F. O'Dea Assistant Examiner-F. D. Shoemaker[57] ABSTRACT A crankshaft for an internal combustion engine comprisinga plurality of crank throws whose crank radii are made to varythroughout the crankshaft revolution. The crankshaft and varying radiicrank throws are interconnected by four gears which rotate an eccentricwhich is placed between the connecting rod and crankpin. Rotation of theeccentric may be clockwise or counterclockwise depending on thearrangement of gears used to generate either epicycloidal orhypocycloidal motion in which the cuspated motion is attenuated byincorporating a pair of non-circular gears which also rotate theeccentric upon the crankpin in a manner which causes the piston to movein the cylinder to a precalculated position during the combustionindependent of the uniform speed of the crankpin.

2Clairm,2DrnwlngF1gures Patented Aug. 29, 1972 3,686,972

2 Sheets-Sheet. 1

INVENTOR.

EDWARD M. MC WHOTIEF 1 1. [/L v/ WW/wfu Patented Aug. 29, 1972 3,686,972

- 2 Sheets-Sheet 2 INVENTOR. EDWARD M. MC WHORTER INTERNAL COMBUSTIONENGINE VARIABLE THROW CRANKSI-IAFI BACKGROUND OF THE INVENTION 1. Fieldof Invention This invention relates to the field of crank mechanisms andin particular to their use in the piston driven internal combustionengine.

2. Prior Art In general the amount of crank torque generated by anyreciprocating piston driven internal combustion engine is proportionalto the magnitude of the tangential force vector acting on the crankradius and is in essence simply the rotative effort applied to thecrankpin. This rotative effort is equal to the system net effort whichis acting on the piston in the direction of the piston axial centerlineand resolved into a force acting along the length of the connecting rodand the normal force acting on the piston wrist pin. At the crankpin,the force acting along the connecting rod is resolved into radial andtangential force components. This latter force acting over the length ofthe crank radius produces the torque delivered by the crankshaft. I havediscovered that the efficiency of this type of system can be greatlyincreased by causing the crank radius to vary in length in a cycloidalmanner synchronized to the rotation of the crankpin in a more effectivemanner.

The properties of cycloidal crank mechanisms are well defined in theliterature under the general heading of kimematics. It should be pointedout with certainty that the invention presented in this disclosure isnot a true cycloidal crank as those defined in the literature since thecrankpin axial center line subscribes a circular path about the maincrank journal which itself remains stationary relative to its rotativecenter. The cycloidal motion imparted to the connecting rod at the crankend is caused by placing an intervening cam between the crank pin andthe connecting rod bearing. This cam mechanism may be made to rotateabout the crankpin axial center line by any convenient manner which willproduce a predictable crank and rod position corresponding to anyposition of the piston during the operating cycle of the engine.

SUMMARY OF THE INVENTION The invention is a crankshaft to be used in theinternal combustion engine. The crankshaft comprises a plurality ofcycloidal 2 gear systems in which the stationary gear of each system isfixedly positioned over the centroid of revolution of the crankshafthaving a plurality of crank throws. The moving or running gear of eachcycloidal gear system is fixedly attached to a shaft which is rotatablymounted in one of the pair of crank arms comprising each crank throw. Tothe other end of the shaft is a fixedly attached non-circular gear. Thenon-circular gear is enmeshed with its duplicate or conjugatenon-circular gear which is fixedly attached to an eccentric which isrotatably mounted between the piston connecting rod and the crankpin ofeach crankthrow. Revolution of the crankshaft induces a rotation of therunning cycloidal gear which is enmeshed in the stationary gear andtransmits this rotation through the connecting shaft to the non-circulargear which is enmeshed and rotates its duplicate or conjugatenon-circular gear attached to the eccentric. Rotation of the eccentricis not uniform relative to the revolution of the crankshaft and is notintended to be circumstantial within the limits of the mechanicalboundary of the system. Instead, conditions more favorable to thecombustion are defined and the piston position is precalculated forthese conditions and approximated by designing the ratio of diameters ofthe non-circular gears to generate a movement of the eccentricapproximating the precalculated piston position using rotational inputfrom the cycloidal running gear. The general effect is to shape themechanical boundaries of piston displacement in accordance withcombustion requirements independent of the circumstantial conditions ofpure mechanical movement. The synergism of circular cycloidal andnon-circular gear motion imparted to components within the system has asits basis of operation the effect of varying the otherwise uniformrotation of the eccentric and uniform speed of the crankpin byperiodically increasing and decreasing the throw of the crankshaft inaccordance with the combustion requirements of the particular enginesystem.

As an example of how this system would operate let us first consider theDiesel cycle. The fuel injection rate of a Diesel engine is limited bythe crank rise rate (R.R.) which effects piston speed, and by theextreme rise of pressure (ERP). As a general engineering rule of thumbthe rise rate (RR) multiplied by the extreme rise in pressure ERP)divided by l0,000 should be 2 or less to prevent combustion knocking.

Eq. (I) ERP X RR X 10 =2 or less for good engine operation.

Where: ERP Peak pressure compression pressure RR ERP/degrees of crankrotation In the standard Diesel engine the piston moves at varying speedwhile the crankpin revolves at uniform speed. The placement of theeccentric between the piston connecting rod and crankpin of such asystem and the rotation of said eccentric imparts a second source ofmotion to the piston which augments or subtracts from that motionimparted by the crankpin. The extreme pressure rise (ERP) can thereforebe partially controlled by the motion imparted to the piston by therotation of said eccentric during the period of fuel injection. When theinitial downward motion of the piston is accelerated by the rotation ofthe said eccentric the fuel injection can be completed earlier duringthe power stroke since the increased combustion volume can more readilyaccommodate the extreme pressure rise (ERP) within the acceptable limitsas designated by Equation (l above. Because of the limiting factor ofthe extreme pressure rise (ERP) on the injection rate, the tendency ofmost Deisel equipment operators is to extend the period of fuelinjection in order to obtain the required power output from theirequipment. This practice results in the decrease in residence time ofthe burning fuel mixture last injected into the combustion chamber whichincreases the rate of unburned hydrocarbons in the exhaust emissions.The present invention makes such a practice unnecessary since the rateof fuel injection can be increased allowing an earlier fuel cut offduring the power stroke.

The present invention may also be beneficially employed within the Ottocycle principle. Considering an electrical spark ignition of a normallyaspirated gasoline and air charge it can be shown that the rotation ofsaid eccentric will allow the combustion reaction to proceed atuniformly higher system pressure and density which tend to push thereaction upward along the Hugoniot curve within the weak detonationrange. This is accomplished by decreasing the initial downward motion ofthe piston by the upward rotation of the eccentric thus producing highereffective combustion pressures for a corresponding greater effectivecrank angle. Operation of the system in this manner provides a morecomplete early combustion thus assuring a decrease in the percent ofemission of unburned hydrocarbons in the exhaust stream. The improvedperformance as indicated by the higher combustion pressures for agreater effective crank angle may be traded off against a decrease infuel to air mixture ratio and lower octane fuels to achieve a lowercombustion temperature with a corresponding decrease in the generationof nitrogen oxides and the use of lead additives which are alsoundesirable products of the exhaust emission.

It is therefore the object of the present invention to provide in amanner hereinafter set forth a crankshaft of the aforementionedcharacter comprising a means of controlling piston speed within somemeasure relative to the uniform speed of the crankpin allowingcombustion to proceed in a manner which will be beneficial in thereduction of harmful exhaust emissions.

It is another object of the present invention to provide a variabilitywithin the crankshaft system described to allow its use in engines whichoperate on the Diesel cycle principle or on the Otto cycle principlewithout loss of thermodynamic efficiency within either system.

It is yet another object of the present invention to provide a means ofdecreasing the size of the associated gear train components by a uniquemethod of constructing the crank arms.

Various other possible objects and advantages of the present inventionwill become apparent to those skilled in the art from the description tofollow which discusses in detail the particular preferred embodimentsand should not be taken as limiting the true scope of the invention asset forth in the appended claims.

BRIEF DESCRIPTION OF THE DRAWING There are included as part of thespecification drawings illustrative of the invention and brieflydescribed as follows:

FIG. 1 is a perspective view of a section of a crank mechanism having aplurality of such sections and showing in cut-away the generation ofepicycloidal motion and the reversal and transmittal of said motion to arotating eccentric mechanism placed upon the crank pm.

FIG. 2 is a section view of a crank mechanism showing the generation ofa hypocycloidal motion and the reversal and transmittal of said motionto a rotating eccentric mechanism placed upon the crankpin.

DESCRIPTION OF THE PREFERRED EMBODIMENT Considering first the mechanicaloperation of the system herein described and referring to FIG. 1. Gear 1is centrally positioned over shaft 13 and firmly affixed to the engineblock 17. Shaft 13 is axially aligned with a plurality of equally spacedsimilar shafts constituting the main bearing shafts of a single ormultiple throw crankshaft which is journaled in the engine block 17.Arms 8 and 9 are fixedly attached to their respective adjacent shaft 13.Pin 7 is fixedly attached to arms 8 and 9 providing a solid link fortransmitting a unison of motion to arms 8 and 9 as they revolve withshaft 13. Gear 2 is fixedly attached to one end of shaft 3 which isrotatably mounted in arm 8. Non-circular gear 4 is fixedly attached tothe other end of shaft 3. Non-circular gear 5 is fixedly attached toeccentric 6. Non-circular gear 5 and eccentric 6 are rotatably mountedon pin 7. Gear 2 is enmeshed with gear 1 and is caused to rotate whenarms 8 and 9 revolve clockwise with shaft 13. The rotary motion inducedin gear 2 is transmitted to non circular gear 4 by connecting shaft 3.Non-circular gear 4 is enmeshed with its mating duplicate non-circulargear 5 causing non-circular gear 5 to rotate in a counter clockwisedirection or opposite to the clockwise direction of the shaft 13. Whenthe crown of the eccentric 6 is pointed downward and the piston 19 is attop center position the counter clockwise rotation of eccentric 6 whichis affixed to non-circular gear 5 increases the initial downward travelof the piston making such a system suitable for operation of the Dieselcycle principle for those reasons previously set forth. Split sleevebearings 11 and 12 fit on the inside and outside surfaces of eccentric 6respectively. Bolt 14 secures the bearing cap to connecting rod 10.Extension 15 is a protrusion on arm 8 which accommodates sleeve bearing16 and shaft 3. The position of extension 15 on arm 8 is determined bythe relative size of gears l and 2. Referring to FIG. 2 gear 2] is aninternally toothed gear which is centrally positioned over shaft 33 andfirmly affixed to engine block 34 so as to be stationary relative to therevolution of shaft 33. Shaft 33 is axially aligned with a plurality ofequally spaced similar shafts constituting the main bearing shafts of asingle or multiple throw crank shaft which is journaled in main bearings35 in the engine block 34. Arms 28 and 29 are fixedly attached to theirrespective adjacent shaft 33. Pin 27 is fixedly attached to arms 28 and29 providing a solid link for transmitting a unison of motion to arms 28and 29 as they revolve with shaft 33. Gear 22 is rotatably mounted onarm 28 and fixedly attached to one end of shaft 23. Non-circular gear 24is fixedly attached to the other end of shaft 23. Non-circular gear 25is fixedly attached to eccentric 26. Non-circular gear 25 and eccentric26 are rotatably mounted on pin 27. Gear 22 is enmeshed in gear 21 andis caused to rotate when arms 28 and 29 revolve with shaft 33 in aclockwise direction. The rotary motion induced in gear 22 is transmittedto non-circular gear 24 by connecting shaft 23. Non-circular gear 24 isenmeshed with its matching duplicate non-circular gear 25 causingnoncircular gear 25 to rotate in the same clockwise direction as shaft33. When the crown of eccentric 26 is pointed downward when the piston39 is at top center position the clockwise rotation of eccentric 26which is affixed to non-circular gear 25 decreases the initial downwardtravel of the piston 39 making such a system suitable for operation onthe Otto cycle principle for those reasons previously set forth. Splitsleeve bearings 31 and 32 fit on the inside and outside surfaces ofeccentric 26 respectively. Arm 28 contains an instep section 36 whichallows the significant reduction in diameter of gear 2] and theconsequent reduction in size of all other gears within the system.Connecting rod 40 is rotatably mounted on sleeve bearing 32 at one endand connected to piston 39 at the other end.

The cycloidal motion imparted to the connecting rod by the rotation ofthe crankpin cam mechanism about the crank greatly increases the torqueof the engine at certain periods of the power stroke. In order tofacilitate the demonstration of this fact the following standard systemdimensions and operating conditions will be maintained throughout thediscussion.

System Dimensions Piston diameter 4.00 inches Connecting rod length 9.60inches Normal crank radius 2.40 inches Crank cam eccentricity 0.50inches System Operating Conditions Compression ratio 7.5 Augmentationfactor 20.8% Combustion temperature 5000R Combustion Pressure 957 psiaSpecific heat ratio L3 Three assumptions are made concerning thecalculations used in this discussion, but are not to be taken in anymanner as imposing limitations on the present invention.

l. The number of revolutions of the crankpin cam mechanism about thecrankpin for each complete revolution of the crankshaft shall be in thesame ratio as the diameter of the stationary gear to the diameter of thegear attached to the cam mechanism.

2. The top of the cam of the crankpin cam mechanism will be pointed inthe downward attitude so as to present the shortest possible crankradius when the connecting rod is centered directly on the piston axialcenterline.

3. Expansion of the combustion gases above the piston shall beisentropic.

The crank radius for the hypocycloidal augmented systems is given by theexpression:

Equation 1 The crank radius for the epicycloidal augmented systerns isgiven by the expression:

where: Equation 2 Rt Augmented crank radius Rc normal crank radius (2.4inches) Ra Cam eccentricity (0.5 inches) 0 Stationary gear diameter bCrankpin cam mechanism gear diameter 0 Crank station angle.

The crank radius for the various crank angle stations for a power strokefrom 0' to 180 for the nodal systems, n, 2n, and 3n for thehypocycloidal and epicycloidal rotational systems, abbreviated as H andE respectively, is presented in Table 1 below.

TABLE I (runk Crank rudius Inches Station n In 3n Angle H E H E H E 0 l900 L900 L900 L900 20 l 915 L957 L960 2.077]

40 l 960 2.096 2.l34 2.506 2.534 2.807 60 2.034 2.266 2.400 2.703 2.9002.9000 2134 2.363 2.726 2.877 2.534 2.69! 2.193 2.518 2.900 2.900 2.2822.339 2.257 2.593 2.727 2.877 2.077 2 203 2.400 2.723 2.400 2.703 L900L900 2.558 2.820 2.!34 2.365 2.077 2.203

I60 2.7262880 L960 2.534 2.69l2 I80 2.9002900 L900 2.900 2.900

It will be seen in Table I that the total stroke of the I-l-2n system isonly 3.8 while the standard unaugmented system would be 4.8 inches. Theclearance volume of the H-2n system is 7.36 inches as compared to the9.28 inches of the standard system. Since there are less hot gases tocontend with in the H2n system the expansion of these gases during thesuction stroke, under real engine conditions would be less. Also theamount of fluid friction losses or throttling around the intake valvewould be less. Both conditions would tend to improve the volumetricefiiciency of this system over the standard system shown.

The relative piston travel corresponding to those crank radius valuesshown in Table l are presented in Table ll together with values of thestandard engine system which shows a normal crank radius of 2.4 inchesfor comparison.

TABLE."

Crank Piston Travel Station Standard n 2n 3n Angle Engine H E H E H E 0O 0 00 0 I0 .048 .050 .01] .042 .034 20 .183 .I73 .013 .283 .l56 40 .696.65] .l40 .693 0 .5l9 0' 60 L440 L355 .538 L406 .432 .784 L02 80 2.2952J6? L252 2.222 L657 L270 2.245 90 2.720 2.576 L727 2348 2.348 L698 2667I00 3.l30 2.969 2.229 2.365 2.978 2.134 2.928 I20 3.840 3.673 3.2372.9l8 3.666 2.992 2.992 [40 4.370 4.220 4.077 3.392 4.05l 3.754 3.338l60 4.690 4.600 4.207 3.695 3.942 4.420 4.339

From Table II it can be seen that the downward piston travel for theE-2n and E-3n systems does not occur until after the crank station angleof 40 is attained. This is caused by the augmentation cam which does notallow the piston to reach the top-deadcenter position until an effectivecrank angle of 3S5l' and 4 l-29' are achieved for each systemrespectively. This means that torque will be generated at thetop-deadcenter position at ignition for these systems which is notpossible with standard crank systems. This condition becomes morepronounced as the augmentation factor is increased.

The approximate cylinder pressures corresponding to the piston travel ofthe various systems in Table ll are presented in Table III.

TABLE lll Crank Cylinder Pressure psia Station Standard n 2n 3n AngleEngine H E H E H E The generally higher cylinder pressures of theaugmented systems compared to those of the unaugmented standard enginesystem, as indicated in Table 111, is caused by the upward movement ofthe cam mechanism during the initial downward movement of the crankwhich shortens the piston travel as shown in Table 11.

The rotation of the cam mechanism about the crankpin causes the crankradius to change in a manner which produces an effective crank anglewhich is different than the crank station angles shown in the abovetables. The adjusted effective crank angles necessary for thecomputation of crank torque is presented in Table 1V.

TABLE IV crank Effective Crank Angle Degrees/Minutes station standard n2n 3n angle engine H E H E H E The angle formed by the connecting rodand piston axial centerline corresponding to the effective crank anglesvalues shown in Table IV are presented in Table V.

TABLE V crank Connecting Rod Angle station standard It 2n 3n angleengine H E H E H E The Crank torque calculated for each crank stationangle using values from Tables 1, 111, IV and V are presented in TableVL.

TABLE V1 crank Crank Torque, lnch/ Pounds station standard n 2n 3n angleengine H E H E H E Torque Ry[sin 0+3) p cos [i where:

T= Torque, inch pounds Rt= Crank radius (Table l) P= Cylinder pressure(Table 111) A= Piston area 12.56 in.

3 Rod angle (Table V) 'y= Effective crank angle (Table IV) Since theinitial top-dead-center positions of the E-2n and the E-3n systems donot occur at the 40 crank station angle the starting values of thesesystems shown in the above tables have been adjusted to reflect theapproximate piston top-dead-center position and are therefore markedwith an asterisk.

The values shown in Table 1 through V1 are approximate and are to thebest knowledge of the inventor correct within the normal range of sliderule accuracy. Pressure values presented in Table 111 were taken from agraphical plot of cylinder volume versus cylinder pressure for anisentropic expanding system and are thought to be accurate within therange of i 5 psi. The values presented are used for illustrative purposeonly in order to describe the unique and beneficial effects of thesystem. Actual point design will adhere to more rigorous methods ofanalysis but will not significantly change the values within scope ofthe illustrative example presented.

The generally higher torque values of the augmented systems shown inTable V1 reflect the beneficial combinations of, longer crank radius,higher corresponding cylinder pressures, more effective combinations ofconnecting rod and effective crank angles, which develope highertangential force vectors at the crankpin.

What is claimed is:

1. in engines of the character described,

a. An engine block,

b. A plurality of equally spaced and axially aligned main journalbearings affixed to said engine block,

c. a plurality of stationary gears centrally positioned over the axialcenter-line of said main journal bearings and being affixed to saidengine block,

d. a crankshaft rotatably mounted in said main journal bearings,

e. said crankshaft having a plurality of shafts and a plurality of armsfixed at the (end) ends of said shafts,

f. said arms being radially aligned in respective pairs and each saidpair affixed in this position by a (common) pin, said pin being firmlyaffixed to each said arm, in a respective pair,

g. an eccentric rotatably mounted on each said pin,

h. a connecting rod rotatably mounted upon each said eccentric saidconnecting rod being connected to a piston at its opposite end,

i. a non-circular gear attached to each said eccentric,

j. each said non-circular gear being enmeshed with (its) a duplicatenon-circular gear,

k. each said duplicate non-circular gear being rotated by a fixedlyattached connecting shaft rotatably mounted on an extension protrudingfrom the edge of one of the said arms comprising each of the said pairof radially aligned arms,

I. the opposite end of each said connecting shaft being fixedly attachedto and rotated by a running gear enmeshed in one of said stationary(gear) gears,

m. each said enmeshed running (gear) gears being caused to turn by therevolution of said crankshaft,

n. said rotation causing each said eccentric to be rotated in theopposite clockwise direction as the rotation of said crankshaft.

2. ln engines of the character described,

a. an engine block,

b. a plurality of equally spaced and axially aligned main journalbearings affixed to said engine block,

c. a plurality of stationary internal tooth annular gears centrallypositioned over the axial center line of said main journals and fixedlyattached to said engine block at a point approximately one third thedistance of the working span between said main journal bearings,

d. a crankshaft rotatably mounted in said main journal bearings,

e. said crankshaft having a plurality of shafts and a plurality of armsfixed at the (end) ends of said shafts,

f. said arms being radially aligned in respective pairs and each saidpair affixed in this position by a (common) pin, said pin being fixedlyattached to each said arm in a respective pair, one arm of each saidradially aligned pair of arms (contain) containing an instep toaccommodate clearance of said stationary internal tooth annular gears,

g. an eccentric rotatably mounted on each said pin,

h. a connecting rod rotatably mounted upon each said eccentric, eachsaid connecting rod being connected to a piston at its opposite end,

i. a non-circular gear attached to each said eccentric, each saidnon-circular gear enmeshed with a duplicate noncircular gear,

j. each said duplicate non-circular gear being rotated by a fixedlyattached connecting shaft each said shaft being rotatably mounted on thesaid arm of each said radially aligned pair of arms containing saidinstep,

k. the opposite end of each said connecting shaft being fixedly attachedto and rotated by a running gear enmeshed in a respective saidstationary internal tooth (gears) gear, I. each said enmeshed runninggear bemg caused to turn by the revolution of said crankshaft,

m. said revolution causing each said eccentric to be rotated in the sameclockwise direction as the revolution of said crankshaft.

I I l

1. In engines of the character described, a. An engine block, b. Aplurality of equally spaced and axially aligned main journal bearingsaffixed to said engine block, c. a plurality of stationary gearscentrally positioned over the axial center-line of said main journalbearings and being affixed to said engine block, d. a crankshaftrotatably mounted in said main journal bearings, e. said crankshafthaving a plurality of shafts and a plurality of arms fixed at the (end)ends of said shafts, f. said arms being radially aligned in respectivepairs and each said pair affixed in this position by a (common) piN,said pin being firmly affixed to each said arm, in a respective pair, g.an eccentric rotatably mounted on each said pin, h. a connecting rodrotatably mounted upon each said eccentric said connecting rod beingconnected to a piston at its opposite end, i. a non-circular gearattached to each said eccentric, j. each said non-circular gear beingenmeshed with (its) a duplicate non-circular gear, k. each saidduplicate non-circular gear being rotated by a fixedly attachedconnecting shaft rotatably mounted on an extension protruding from theedge of one of the said arms comprising each of the said pair ofradially aligned arms, l. the opposite end of each said connecting shaftbeing fixedly attached to and rotated by a running gear enmeshed in oneof said stationary (gear) gears, m. each said enmeshed running (gear)gears being caused to turn by the revolution of said crankshaft, n. saidrotation causing each said eccentric to be rotated in the oppositeclockwise direction as the rotation of said crankshaft.
 2. In engines ofthe character described, a. an engine block, b. a plurality of equallyspaced and axially aligned main journal bearings affixed to said engineblock, c. a plurality of stationary internal tooth annular gearscentrally positioned over the axial center line of said main journalsand fixedly attached to said engine block at a point approximately onethird the distance of the working span between said main journalbearings, d. a crankshaft rotatably mounted in said main journalbearings, e. said crankshaft having a plurality of shafts and aplurality of arms fixed at the (end) ends of said shafts, f. said armsbeing radially aligned in respective pairs and each said pair affixed inthis position by a (common) pin, said pin being fixedly attached to eachsaid arm in a respective pair, one arm of each said radially alignedpair of arms (contain) containing an instep to accommodate clearance ofsaid stationary internal tooth annular gears, g. an eccentric rotatablymounted on each said pin, h. a connecting rod rotatably mounted uponeach said eccentric, each said connecting rod being connected to apiston at its opposite end, i. a non-circular gear attached to each saideccentric, each said non-circular gear enmeshed with a duplicatenon-circular gear, j. each said duplicate non-circular gear beingrotated by a fixedly attached connecting shaft each said shaft beingrotatably mounted on the said arm of each said radially aligned pair ofarms containing said instep, k. the opposite end of each said connectingshaft being fixedly attached to and rotated by a running gear enmeshedin a respective said stationary internal tooth (gears) gear, l. eachsaid enmeshed running gear being caused to turn by the revolution ofsaid crankshaft, m. said revolution causing each said eccentric to berotated in the same clockwise direction as the revolution of saidcrankshaft.